Power transmission apparatus

ABSTRACT

Power transmission apparatus including conventional means for enabling manual or automatic shifting of ratios from a suitable low gear ratio to a direct drive ratio and an internal overdrive means for enabling the drive ratio to vary automatically in accordance with the vehicle load so that the relationship between vehicle load and engine torque is maintained in the most favorable condition at all points in the cruise range. This is accomplished by means of an additional reaction gear train which automatically varies the overdrive ratio in accordance with the applied load.

United States Patent [1 1 Herzog Sept. 17, 1974 lnventorl William 13800 Almade" 1,334,021 6/1963 France 74/688 Rd., San Jose, Calif. 95120 [22] Filed: Aug. 18, 1972 Primary ExaminerBenjamin W. Wyche Assistant Examiner-P. S. Lall [21] Appl 281912 Attorney, Agent, or FirmSchatzel & Hamrick Related US. Application Data [63] Continuation-in-part of Ser. No. 77,594, Oct. 2,

1970, abandoned. [57] ABSTRACT Power transmission apparatus including conventional [52] US. Cl. 74/682, 74/688 means for enabling manual or automatic shifting of ra- [51] Int. Cl. F16h 37/06, Fl6h 57/10 tios from a suitable low gear ratio to a direct drive [58] Field of Search 74/682, 688, 785 ratio'and an internal overdrive means for enabling the I drive ratio to vary automatically in accordance with [56] References Citedthe vehicle load so that the relationship between vehi- UNITED STATES PATENTS- cle load and engine torque is maintained in the most I favorable condition at all points in the cruise range. ga 2: This is accomplished by means of an additional reac- 3:242:769 3/1966 Johnson"; 74/682 tion gear train which automatically varies the over- 3,299,741 1/l967 Twiford 74/682 drive ratio in accordance with the pp load- 3,468,l92 9 1969 N 74 3,500,704 1970 asvyns [682 4 Claims, 13'Drawing Figures Muller et al. 74/688 LOCATION OF CENTERLINE OF'GEARS IN 3 PAIENIEDSEP I 11914 v 3.835.731

sum 1 OF 9 LOCATION OF CENTERLINE OF GEARS m Flea IN V EN TOR WILL/AM s. HERZOG PAIENIED I Baas-1.131

SHLET 2 BF 9 INVENTOR WILL/AM S HERZOG PAIENTEUsEP 1 71914 SHEU 3 m 9 '%OF GEARS IN VEN TOR PMENIEBSEPY'Q" 3.935.731

suwumg v IN V EN TOR WILLIAM S: HEl-PZOG PAIENTEBSEP I 11914 SHEET 6 BF 9 INVENTOR WILL/AM .SI HERZOG POWER TRANSMISSION APPARATUS This application is a continuation-in-part of my previous application entitled Transmission, Ser. No. 77,594 and filed Oct. 2, 1970, now abandoned.

BACKGROUND OF THE INVENTION The present invention relates generally to power transmission apparatus and more particularly to an improved transmission for automotive application.

As is well-known by those skilled in the art, the vast majority of all driving occurs with the transmission operating in the direct drive range. Since various types of engines have particular non-linear engine speed/engine torque characteristics, axel gearing ratios must be chosen so that the most suitable portion of the speed/- torque characteristic corresponds to the vehicle power requirements over the intended operational range. Quite obviously, the vehicle powerrequirements over its speed range are also non-linear and there is typically no correspondence between the engine torque nonlinearities and the vehicle speed/power requirement non-linearities.

For example, most automotive drive train designs are chosen so that the most efficient operational speed is about 45 miles per hour. Quite clearly, efficiency falls off on either side of this speed. Since most freeway driving today is done at speeds more than miles per hour faster than the most efficient vehicle speed, it should be quite apparent that efficiency could be vastly improved by providing a variable overdrive feature which enables the engine torque requirements to be automatically matched with the driven load.

SUMMARY OF THE INVENTION It is therefore a primary object of the present invention to provide an automatic transmission having an overdrive feature in which the drive ratio changes in proportion to the applied vehicle load.

Briefly, the preferred embodiment of the transmission includes conventional means for enabling manual or automatic shifting of ratios from a suitable low-gear ratio to a direct-drive ratio, and an internal overdrive means which enables the drive ratio to vary automati cally in accordance with the vehicle load so that the relationship between vehicle load and engine torque is maintained at the most favorable condition at all points in the cruise range. This is accomplished by means of an additional reaction gear train which automatically varies the overdrive ratio in accordance with the applied load.

Among the many advantages of the present invention is that the efficiency of operation of the automobile drive system is substantially improved.

Another advantage of the present invention is that since the engine will be enabled to operate at its most efficient speed, combustion will be made more complete, resulting in less automotive pollution.

These and other advantages of the present invention will no doubt become apparent to those of ordinary skill in the art after having read the following detailed description of the preferred embodiments which are schematically illustrated in the several figures of the drawing.

IN THE DRAWING FIG. 1 is a sectional view taken along the longitudinal axis of a simplified embodiment of a transmission'in accordance with the present invention;

FIG. 2 is a schematic diagram illustrating the interrelationship between various gears of the embodiment illustrated in FIG. 1;

FIG. 3 is a sectional view of supplementary gears located in the position noted in FIG. 1;

FIG. 4 is a diagram illustrating the relationship between the gears located along line 4-4 of the FIG. 1 embodiment;

FIG. 5 is a diagram illustrating an alternative gearing configuration in accordance with the present invention;

FIG. 6 is a sectional view taken along the longitudinal axis of an alternative embodiment of the present invention;

FIG. 7 is a sectional view further illustrating the relationship between the gears shown at 7 in FIG. 6;

FIG. 8 is a diagram illustrating the relationship between the gears located along the line 8-8 of FIG. 6;

FIG. 9 is a sectional view taken along the longitudinal axis of another alternative embodiment of a transmission in accordance with the present invention;

FIG. 10 is a sectional view taken along the lines l010 in FIG. 9;

FIG. 11 is a diagram illustrating other gears shown in part in the embodiment of FIG. 9;

FIG. 12 is a sectional view taken along the longitudinal axis of still another alternative embodiemnt of a transmission in accordance with the present invention;

FIG. 13 is a diagram illustrating the relationship between certain gears appearing in the embodiment illustrated in FIG. 12.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT Referring now to FIG. 1 of the drawing, a simplified embodiment of a basic variable drive ratio transmission in accordance withthe present invention is indicated generally by the reference number 15. The transmission is enclosed in a housing 16 which is mated to the rear-end of a driving engine by the bell-housing 17. The engine crankshaft 18 and flywheel 19 are secured to an adaptor plate 20 for driving the sun-gear 22 of an input planetary system 21. In addition to sun-gear 22, input planetary gear system 21 includes a plurality of planetary pinions 23 equally spaced around sun-gear 22 and in meshing relationship therewith. Each planetary pinion 23 is mounted on a stub shaft 24 of a spider 25.

The input system further includes a ring-gear 26 which meshes with the planetary pinions 23 and is held in place by the two spaced apart flanges 27 of a retainer 28. A brake band 29 is disposed around the outer periphery of ring-gear 26 for permitting rotation of ringgear 26 to be selectively prevented. As illustrated, it will be noted that the end 30 of an input shaft 31 is fitted into a flanged bearing 32 located in the center of sun-gear 22.

The hub 33 of spider 25 is afiixed to input shaft 31 by suitable means (not shown) and is journaled to the bearing support plate 35 by a ball-bearing assembly 34. Retainer 28 is also supported by plate 35. I-Iub 33 provides a means for rotating shaft 31.

Disposed concentric with input shaft 31 and journaled thereto by a pair of bearings 36 is a cylindrical hub 37 which carries a brake drum 38 at one end and a spur gear 39 at the other end. Disposed about brake drum 38 is a brake band 39 which may be actuated by conventional means to impose a selected drag on the brake drum 38 for purposes which will be explained below.

Also disposed about input shaft 31 is a regulator assembly 45 which includes a hub 47 that is journaled to housing 16 by the bearing assemblies 471 and 472. Hub 47 carries an internal gear 44 having gear teeth on its inner periphery for engaging a gear 43 which drives a shaft 401 that is journaled to housing 16 by a pair of bearings 402 and 403. A second gear 40 is affixed to shaft 401 and meshes with a gear 41 shown in FIG. 3. Referring now briefly to FIG. 3, it will be noted that gear 41 is journaled to a shaft 411 by a suitable bearing means 412 and serves to drive the gear 42. Shaft 411 is non-rotatably mounted to housing 16. Gear 42 meshes with spur gear 39.

As illustrated in FIG. 1, the leftmost end of hub 47 also carries the ring-gear 48 of an output planetary gear system 480 tha further includes a plurality of planetary pinions 49 and a sun-gear 52. The planetary pinions 49 are carried by the stub shafts 50 of a spider 51 which is keyed to the leftmost end of input shaft 31, and mesh with ring-gear 48 and sun-gear 52. The end 54 of input shaft 31 extends into a bearing equipped opening shaft 53 which is journaled to housing 16 by a ball-bearing assembly 56.

In FIG. of the drawing, the inter-relationships be tween the various gears forming the regulator assembly 45 is shown.

In operation with brake band 29 applied, and with control brake band 39' released, sun-gear 22 drives pinions 26 which in turn drive spider 25 to drive input shaft 31. Since no drag is applied to drum 38, the gears of regulator assembly 45 will be free to revolve and thus ring-gear 48 will be free to revolve. Accordingly, the planetary pinions 49, when driven by input shaft 31 via spider 51, will merely revolve around sun-gear 52 and no drive will be imparted to output shaft 53. However, if then a braking force is applied to drum 38 by brake band 39, revolution of ring-gear 48 will be slowed causing sun-gear'52 to be rotated by pinions 49. Thus, it will be appreciated that the drive ratio between input shaft 31 and output shaft 53 will be directly related to the braking force applied to drum 38.

As shown in the embodiment illustrated in FIG. 1, (see also FIG. 4) the proportional size of the gears of the input planetary gear system have a ratio of 3:1 and are consequently capable of increasing the engine torque approximately three times.

In operation with brake band 29 released, it will be appreciated that the transmission is in the neutral configuration and the engine is permitted to turn sun-gear 22 freely without any drive energy being transmitted to the output shaft 53. However, when brake band 29 is actuated to lock ring-gear 26 in place, rotation will be imparted to shaft 31 through the planet pinions 23 and spider 25. As a result, the planet pinions 49 and the output planetary system will be rotated since spider 51 is affixed to the output end of shaft 31 and in rotating will cause output shaft 53 to be driven at a speed determined by the rotational differential between shaft 31 and ring-gear 48 by slowing the speed of ring-gear 48 to less than that of input shaft 31. It will be appreciated that the speed of output shaft 53 will thus be increased.

A selected drag is applied to ring-gear 48 as a result of the braking force applied to brake drum 38 by the brake band 391.

More particularly, note that as ring-gear 48 turns, it also turns ring-gear 44 which in turn drives gear 43. Gear 43 drives gear 40 which in turn drives gear 41 and gear 42. Gear 42 meshes with gear 39 which in turn drives brake drum 38. Accordingly, it will be seen that with no braking force applied to drum 38, ring-gear 48 will be free to rotate with shaft 31 (ignoring of course the frictional drag in the regulator gears), and the output shaft 53 will rotate at the same speed that the shaft 31 rotates so long as the load is small. However, if the load on output shaft 53 is materially increased, regulator 45 will permit shaft 53 to turn at a lesser speed than shaft 31. However, by applying brake band 391 to provide a drag force to brake drum 38, the effect will be to slow the rotation of ring-gear 48 to less than that of shaft 31 and accordingly, sun-gear 52 and output shaft 53 will be driven at a higher speed than is shaft 31. This of course, is the overdrive configuration. The proportion of the gears in the output planetary system are chosen to have a ratio of 1:5 of the input shaft to the output sun-gear 52 and to the output shaft 53 to which it is secured. If brake drum 38 is brought to a halt, then the ratio of the engine speed to that of the output shaft 53 will be 1:1 Accordingly, the drive ratio can be selectively varied from 1:0 to 1:1 These ratios are of course for illustrative purposes only and may or may not be used in practice. While some energy is lost due to frictional losses in the control brake band, this may not be excessive for some application.

Turning now to FIG. 6 of the drawing, an alternative embodiment of the invention is shown in simplified diagrammatic form, eliminating non-essential structure such as housing bolts, etc. In this embodiment, the end 18 of the crankshaft of an internal combustion engine is suitably secured to flywheel l9 and as in the previous embodiment, an adaptor plate 20 is secured to flywheel 19. The sun-shaft 61 of a sun-gear 62 is secured to plate 20 by means of a splined flange. Sun-gear 62 forms part of an input planetary gear system 21 which additionally includes a plurality of planetary pinions 67 and a ringgear 58. A brake band 59 is disposed about the outer periphery of ring-gear 58 for allowing rotation thereof to be selectively terminated.

Ring-gear 58 is provided with an elongated hub 60 that is journaled to rotate freely about the shaft 61 affixed to sun-gear 62. Hub 60 is spline connected to flange 63 which is in turn, secured to a clutch 64 having a clutch-plate 65 as shown in the drawing. A clutch release bearing assembly 66 is provided in the location illustrated Each of the three planetary pinions 67 are rotatably mounted to the stub shaft 68 of a spider 69 which is secured by one or more keys or splines to an input shaft 31 which extends therethrough with its end 30 being rotatably received within the bearing fitted bore 32 of sun-gear 62. The rearmost portion of shaft 31 extends through an axial opening in worm-wheel 83. Worm-wheel 83 is journaled to housing 16 by a pair of bearings and 91. The end 54 of shaft 31 is secured to the spider 51 of an output planetary gear system illustrated generally at 480. Output planetary system 480 includes a ring-gear 48 which is affixed to worm-wheel 83, a plurality of planetary pinions 49 which are carried by the stub shafts 50 of spider 51, and a sun-gear 52 which is affixed to one end of an output shaft 53.

Disposed between the output planetary system 480 and the previously mentioned input planetary system 21 is a regulator assembly (reactor means) generally designated 92, which includes the worm-wheel 83, a worm-gear 84 which drives a pair of gears 85 and 86 (see FIG. 5) that in turn drive a pair of gears 80 and 81 that in turn engage one of the spur gears 78 and 79 carried by an axially movable sleeve 82 (see FIG. 7). Sleeve 82 is spline connected to one end of a hollow shaft 77 that is disposed concentric with input shaft 31. The gears 78 and 79 may be shifted in and out of mesh with gears 80 and 81 by means of a control fork (not shown) which fits into the peripheral recess indicated generally at 7 in sleeve 82.

The opposite end of shaft 77 is operatively coupled to a brake drum 75 having a brake band 76 disposed thereabout for controlling its speed of rotation. As a means for controlling the regulator without undue losses due to friction and to reduce wear on the braking surfaces of brake band 76, a hydrokinetic device illustrated generally at 94 is utilized. At 95 a splined bearing sleeve is provided on which the guide wheel 74 of hydrokinetic device 94 is supported yet permits the splined hollow shaft 77 to rotate when the guide wheel is prevented from rotating by the actuation of a brake band 73 which engages the brake drum 72. Brake drum 82 is secured to guide wheel 74 and is also with 95, supported by a bearing journaled to spider 69. The impel, ler 71 of hydrokinetic device 94 is shown secured to the splined hollow shaft 77. The runner 70 of hydrokinetic device 94 is secured to the spider 60 which is, as indicated above, spline connected to the input shaft 31. With the regulator designed so that the splined hollow shaft 77 rotates in the same direction as the input shaft 31 and with brake band 76 disengaged and brake band 73 actuated, the hydrokinetic device 94 will feed back into the input shaft 31 a portion of the control energy. If it is desired that the regulator be brought to a halt, then the brake band 73 will be released and brake band 76 will be applied.

For reasons of simplicity, the mechanisms for actuating the brake bands 59, 73 and 76 are not shown in any of the views of the drawing. Similarly, while the bearings of the rotating parts are illustrated as either ballbearings or plain bronze bearings, it will be realized that these conventions are chosen for purposes of illustration and are not limiting. Furthermore, the various gears illustrated can be of any suitable configuration. In operation, this embodiment of the present invention is very similar to that described above with regard to the first simplified embodiment, except that instead of using a brake drum and brake combination as the primary means of exerting drag on the ring-gear of the output planetary system 480, the hydrokinetic device 94 is utilized.

When the engine is being started, or the automobile is being towed for service, the lockout clutch 64 is disengaged and all of the brake bands 59, 73 and 76 are released. With the control gear 79 in mesh with gear 81 to prevent excessive rotational speed of the means of resistance to rotation, with the lockout clutch 65 remaining disengaged, with the high torque brake band 59 fully applied, with brake band 73 gradually applied to bring brake drum 72 and guide wheel 74 to a halt, the impeller 71, which is splined to the hollow shaft 77, resists the rotation of the control gears and reactor members of the regulator 83. When the impeller 71 has been slowed down to approximately the same rotational speed'of the input shaft 31, torque is imposed on the output shaft 53 to a degree according to the design of the first and second planetary systems. For example, in the first planetary, or high torque gearing, ring-gear 58, sun-gear 62 and planetary gears 67 have a ratio of 6:1, that is, for six revolutions of the sun shaft 61, the input shaft 31 revolves one time.

In the second planetary gear system, the ring-gear 48 (disregarding the fact that it may be revolving and assuming that it is fixed) the sun-gear 52 and the planetary gears 49 have a ratio of 1:2; that is, there will be two revolutions of the sun-gear for each revolution of the input shaft 31. The torque, but not the speed of the cranksahft 18, relative to the output shaft 53, will be approximately 3:1. When the impeller has been slowed down to approximately the same rotational speed as the input shaft, the engine is revolving rapidly supplying the additional torque that is required. When the brake band 73 is released and brake band 76 is applied, the torque on output shaft 53 remains fairly constant, but the speed of the output shaft 53 increases to the fixed ratio of 3:1. In other words, for three revolutions of the crankshaft 18, there will be one revolution of the output shaft 53. To bring the transmission into the economy range, and the maximum overdrive, the brake band 76 is released, the high torque brake band 59 is also released, the control gear 79 is shifted out of mesh with gear 81, and control gear 78 is shifted into mesh with gear 80 so that the hydrokinetic will be effective at low rotational speeds of the input shaft 31. The lockout clutch is also engaged.

The brake band 73 is then again applied, bringing the rotation of the brake drum 72 and the guide wheel 74 to a halt. When the impeller 71 has again been slowed down to approximately the same rotational speed as the input shaft 31, the brake band 73 is released and the brake band 76 is applied. The control gear train and the reactor members are now stationary and the transmission is in maximum overdrive. The crankshaft of the engine is revolving at the minimum RPM with the engine at full power. Should the automobile begin to ascend a grade requiring more torque, brake band 76 will be allowed to slip to in turn permit the engine to pick up speed, providing more torque to the output shaft 53. If still more torque is needed, the brake band 76 will be released and the brake band 73 will be applied to make the hydrokinetic device operate allowing more slip and still further increase in speed of the engine. This will continue as required until the maximum torque obtainable with the high torque gearing locket out is reached, which would be approximately when the output shaft is revolving at near the same RPM as the crankshaft. Further reduction of the resistance to rotate, causes the torque to fall off rapidly. To obtain more torque if needed the high torque planetary gearing must be made operative.

Hydrokinetic device 94 is adjusted so as to transmit just enough energy to runner to balance the energy fed back through regulator assembly 92 so as to cause ring-gear 48 to rotate at a slower rate than does input shaft 31; the effect is to cause sun-gear 52 to be driven at a rotational speed faster than that of input shaft 31, hence, the overdrive action. However, should the torque requirements of output shaft 53 be increased, i.e., the driven load increase, as in going up a hill, the rotational speed of sun-gear 52 will tend to decrease causing the planetary pinions 49 to drive ring-gear 48 faster feeding back more energy through regulator assembly 92 to the hydrokinetic device 94 to counter the imposed drag force. Device 94 permits this variation and thus allows the overdrive ratio to move in the direction of direct drive, i.e., toward a 1:1 ratio. However, at such time as the increased load tends to decrease, the overdrive ratio will tend to increase again to that ratio permitted by the drag imposed by hydrokinetic device 94.

The hydrokinetic device 94 was chosen for this embodiment as a means of resistance to rotation because it provides a good illustration of what is needed to obtain the desired characteristics, mainly, a means that can slip continuously without injury or wear and one that can be made so that the amount'of resistance to fed back energy can be selectively varied. The hydrokinetic device 94 could of course, be replaced by any one of numerous types of devices that would serve as a means of resistance to rotation either mechanical, electrical, frictional, numatic or otherwise. The means, either automatic or manual, for applying and controlling the various brake bands and the lockout clutch, and for shifting the control gears 78 and 79, are not shown for the sake of simplicity. Any of the sensing electrical and hydraulic mechanisms commonly used on various automatic transmissions, could of course, be adapted for use in the present embodiment. This embodiment is deemed suitable for use in light compact automobiles and other comparable machinery from which average performance is expected. However, with the addition of other gearing, for example, a supplementary planetary system to the first or high torque planetary system, or with the addition of a supplementary planetary system to the second planetary system, or both, the performance could be substantially improved.

Although the above description of regulator 92 has been given in terms of a means for reacting to energy fed back from ring-gear 48, it will be appreciated that the control sequence can also be thought of as a parallel drive system in which a small amount of the input energy is bled off and used to slow the rotational speed of ring-gear 48. Since a differential in the rotational speeds (about the common axis) of planetary pinions 49 and ring-gear 48 will result in sun-gear 52 having a higher rotational speed than either, it will be appreciated that the overdrive function (drive ratio of less than 1:1) is likewise accomplished. This is to say that the torque supplied to output shaft will be equal to the difference between the energies input to planetary pinions 49 and ring-gear 48. Accordingly, since the input torque must always equal the output torque, it will be appreciated that where the output torque requirement increases, there will be a tendency toward greater resistance to the effect of the ring-gear retarding force with the result being to increase the speed of ring-gear 48 toward that of the planetary pinions 49, thus causing the drive ratio to move toward 1:1.

Referring now to FIG. 9 of the drawing, another modification of the present invention is shown. In this embodiment the flywheel 100 is affixed to an adapter plate 102 which is secured to one end of a shaft 104. Secured to the opposite end of shaft 104 is a sun-gear 106 which meshes with three planetary gears 108 that are contained within the ring-gear 110 and form an input planetary system 115. Ring-gear 110 is mounted concentric with shaft 104 and is secured to a lockout clutch plate 112 by a sleeve 111. Plate 112 forms part of a conventional friction clutching mechanism 114 for driving ring-gear 110. Ring-gear may be selectively locked against rotation by a brake band 116. The aforementioned structure is contained within a bellhousing 117 which forms a portion of the transmission housing 118.

The planetary gears 108 drive a spider 120 which is journaled to housing 118 as indicated at 122 and is affixed to one end of a transmission input shaft 124 which extends therethrough and is journaled into the end of sun-gear 106 as shown at 119. A control clutch mechanism 126 including a friction surface 128 and the runner 130 of a fluid clutch is aflixed to the spider 120. The impeller 132 of the fluid clutch is carried by a spider 134 which is mounted coaxially with input shaft 124. Spider 134 carries three planetary gears 136 about a sun-gear 138. Sun-gear 138 forms one end of a cylindrical sleeve 140 that is disposed coaxial with input shaft 124 and is displacable along the axis of shaft 124 by means of a shifting fork 142 which pivots about an axis 144. At the opposite end of sleeve 140, a second gear 146 is formed. Sleeve 140 is also provided with an annular flange 150 having a friction surface 152 for engaging the surface 128 of control clutch mechanism 156. The outer perimeter of flange 150 also forms a friction surface 154 for engagement by a second braking band 156.

Also mounted concentrically with respect to sleeve 140 and input shaft 124 is a cylindrical inner sleeve 158 having a flange 160 provided at one end and means forming a ring-gear 162 provided at the opposite end. Ring-gear 162 is engaged by the planetary gears 136. Sleeve 158 is bearing mounted to turn independent of both shaft 124 and sleeve 140 and is axially displaced along shaft 124 by sleeve 140. Flange 160 carries a friction disc 161 providing a friction surface as experienced below.

An output shaft 164 is connected to the sun-gear 166 of an output planetary system 167. Sun-gear 166 is operatively coupled to input shaft 124 by three planetary gears 168 that are carried by a spider 170 affixed to the end of input shaft 124. Planetary gears 168 also engage a ring-gear 172 having an outer periphery forming a friction surface 174. A third brake band 176 is provided for engaging surface 174 to lock ring-gear 172 relative to housing 118. Ring-gear 172 drives a sleeve 178 which is journaled about shaft 164 by a bearing member 180, and about input shaft 124 by a pair of bearing members 182 and 184. Affixed to sleeve 178 is an annular worm bear 188.

Rotatably mounted concentric with sleeve 178 via the bearing members and 192 is a reactor frame 194 having an outer periphery forming a braking surface 196. A fourth braking band 198 is provided for engaging surface 196 to lock frame 194 relative to housing 118. Frame 194 carries a first set of control gears 200 and 202, which are perhaps better illustrated in the partial section shown in FIG. 10 of the drawing. As indicated, the gears 200 and 202 are mounted on a common shaft 204 which is journaled to frame 194 by a pair of bearings 206 and 208. Gear 200 is a helical gear suited for engaging worm gear 188.

Also carried by frame 194 is a second set of control gears 210 and 212 which are mounted to a shaft 214 that is journaled to frame 194 by bearings 215 and 217 as illustrated in FIG. 1 1. Gear 210 engages the gear 146 and gear 212 engages the gear 202. The wall 216 of frame 194 forms a bearing surface for frictionally engaging the friction disc 161 of flange 160 to effectively couple frame 194 to ring-gear 162 via the sleeve 158.

By actuating the four brake bands and the several clutching mechanisms in various combinations, this embodiment can be made to assume at least six operative configurations. In the first configuration, which is equivalent to the neutral position of most standard automatic transmissions, lockout clutch 112 is disengaged, brake bands 116, 156, 176 and 198 are disengaged, and no force is applied to shifting fork 142; thus no frictional engagement is effected between surfaces 128 and 152 or between surfaces 161 and 216.

With lockout clutch 112 disengaged, it will be apparent that as flywheel 100 is rotated, thereby turning sungear 106, ring-gear 110 will be free to rotate in the opposite direction and thus the planetary gears 108 will remain stationary while rotating about their own axes. Input shaft 124 therefore does not rotate and accordingly, no energy is transmitted to output shaft 164. Similarly, since none of the internal brake bands are actuated, output shaft 164 may be rotated freely without damage to any of the internal components of the transmission. This configuration can thus be used where it is required that the vehicle be towed.

In the second, or locked neutral configuration, which is equivalent to the park setting for standard automatic transmissions, lockout clutch 112 is disengaged, brake bands 116 and 198 are engaged while brake bands 156 and 176 are disengaged, and shifting fork 142 is rotated clockwise causing sleeve 140 to be shifted rightwardly to engage the clutching surfaces 128 and 152.

Tracing the resulting drive train through the transmission structure, it will be noted that there is a primary drive path and a secondary drive path which cancel in the output planetary network so that output shaft 164 remains fixed. The primary drive path may be traced from flywheel 100 through adapter plate 102 and shaft 104 to sun-gear 106, planetary gears 108 and spider 120 to input shaft 124. As input shaft 124 is rotated, it in turn rotates spider 170 causing the planetary gears 168 to be revolved about output sun-gear 162. The secondary drive train may be traced through spider 120, control clutch 126, flange 150, sleeve 140, gear 146, gear 210, shaft 214, gear 212, gear 202, shaft 204, gear 200, worm gear 188 and sleeve 178 to ring-gear 172. However, in tracing this drive train, it will be noted that rotation of ring-gear 172 is in the same direction as is spider 170 and rotates at an angular speed which exactly cancels out the effect of rotation of spider 170. Therefore, with the transmission in this configuration, the prime mover may turn flywheel 100 at any speed, or not turn it at all, and output shaft 164 will remain locked.

In the third configuration, which I choose to refer to as the high torque configuration (equivalent to a low gear setting) lockout clutch 112 is disengaged, brake band 116 is tightened, shifting fork 142 is rotated clockwise to engage the control clutching surfaces 128 and 152, and the remaining brake bands are disengaged. With brake band 116 holding ring-gear 110 fast against rotation, rotation of flywheel 100 will cause sun-gear 106 to rotate planetary gears 108 and in turn, rotate spider 120. Since none of the remaining brake bands are engaged, the various sections of the transmission will be free to rotate as a unit and consequently the output shaft 164 will rotate at a reduced speed dependent only upon the gear ratio chosen for the input planetary system including the planetary gears 108. In the preferred embodiment, this ratio was chosen such that 4 1/2 revolutions of the flywheel are required to cause one revolution of output shaft 164.

In the intermediate drive range, the lockout clutch 112 is disengaged, brake bands 116, 156, 176 and 198 are applied, and the force applied to fork 142 is released to disengage the control clutching surfaces 128 and 152. in this configuration, it will be noted that although input shaft 124 is as in the previous configuration, one revolution for every 4 /2 revolutions of flywheel 100, due to the characteristics of the input planetary system 115, the output planetary system 167 causes this ratio to be approximately halved so that output shaft 164 is driven through only one revolution for each two revolutions of flywheel 100. Note that even though the impeller 132 of the fluid clutch may be driven by runner which causes spider 134 to rotate the planetary gears 136 merely roll around sun-gear 138 since ring-gear 162 is free to turn, i.e., surface 161 and 216 are not in engagement.

In the direct drive configuration, lockout clutch 112 is engaged, and shifting fork 142 is rotated clockwise to engage control clutch surfaces 128 and 152. Brake bands 116 and 156, 176 and 198 however are disengaged. In this configuration the ring-gear 110, planetary gears 108 and sun-gear 106 are driven as a unit as flywheel 100 is rotated thereby causing spider 120, input shaft 124 and the remainder of the internal components of the transmission to rotate as a unit so that output shaft 164 is caused to rotate through one revolution for each revolution of flywheel 100.

The sixth operative configuration is the overdrive configuration wherein lockout clutch 112 is engaged, shifting fork 142 is rotated counterclockwise so that the friction or braking surfaes 161 and 216 are engaged, brake band 198 is actuated to engage frame 194 to lock it in position, and brake bands 1 16, 156 and 176 are all disengaged. As in the previous configuration, input shaft 124 is driven in a one-to-one relationship with flywheel 100 causing sun-gear 166 to be rotated at the same speed so long as ring-gear 172 also turns with it, a variable rate depending upon the amount of slippage permitted to ring-gear 172.

Note that since frame 194 is held stationary by brake band 190, a mechanical feedback path is established between ring-gear 172 and the fluid clutch mechanism. This feedback path is through worm 188, control gears 200 and 202, control gears 212 and 210, gear 146, gear 138 and the planetary gear system including the planetary gears 136 and ring-gear 162. Note that since ringgear 162 is held stationary by the engagmeent of surface and wall surface 216, spider 134 is driven via the planetary gears 136 by sleeve gear 138. The variable speed characteristic is effected by the drag imposed upon the fed back energy by the fluid clutch. This feedback path establishes a variable speed range depending upon load requirements and enables output shaft 164 to be driven at a rotational velocity varying upwardly from the direct drive configuration to a potential maximum with brake band 156 applied 2 .4 revolutions for each revolution of flywheel 100.

More specifically, since rotation of ring-gear 172 is reflected through the secondary gear train to impeller 132 of the fluid clutch and tends to cause rotation of impeller 132 in the same direction as that of runner 130, but faster, it will be appreciated that the rotational speed of output shaft 164 relative to input shaft 124 will be variable over a wide range and will be determined by the load applied to output shaft. In other words, with no load applied to output shaft 164, runner 130 would drive impeller 132 at approximately its same rotational speed, thus causing ringgear 172 to be driven in the same direction as spider 170. This of course causes the rotational velocity of output shaft 164 to rotate considerably faster than input shaft 124 (a definite ratio would be vary difficult to determine due to constantly changing factors and could last only momentarily). However, as the load on output shaft 164 is increased, the load will be reflected back through ring-gear 172 causing slippage in the fluid clutch and at some point will cause the rotational velocity of output shaft 164 to approach that of the input shaft 124.

The fluid clutch mechanism therefore plays the role of not only providing for smooth transition between the various drive ranges, but also serves to provide a drag force for implementing the overdrive function. It will of course be appreciated that instead of the fluid clutching mechanism, any equivalent drag source could be utilized in the feedback path. For example, a torque converter, such as that previously illustrated, or a conventional friction clutch might be utilized.

When the transmission is being operated to the overdrive configuration and the load conditions become such as to tend to overload the engine, a shift down to the direct drive configuration is accomplished very simply by releasing brake band 198 and rotating shifting fork 142 clockwise to disengage the frictional surfaces 160 and 216, and to re-engage the control clutching surfaces 128 and 152.

Although not shown in detail herein, it will be appreciated that conventional transmission shifting and range control apparatus can be used to shift the transmission into and out of the various drive configurations.

Referring now to FIG. 12 of the drawing, still another embodiment of the present invention is shown. In this embodiment the flywheel 300 is affixed to an adaptor plate 302 which is secured to one end of a shaft 304. Secured to the opposite end of shaft 304 is a sun-gear 306 which meshes with three planetary gears 308 that are contained within the ring-gear 310 and form an input planetary system 315. Ring-gear 310 is mounted concentric with shaft 304 and is secured to a lockout clutch 312 by a sleeve portion 311. Clutch 312 forms part of a conventional lockout clutching mechanism for driving ring-gear 310. The clutching mechanism is driven into and out of engagement with flywheel 300 by means of a suitable positioning fork 313 which is actuated by conventional actuating linkages. Ring-rear 310 may be selectively locked against rotation by a brake band 316.

The planetary gears 308 drive a spider 320 which is journaled to housing 318 as indicated at 322 and is affixed to one end of a transmission input shaft 324 which extends therethrough and is journaled into the end of sun-gear 306 as shown at 319. interposed between the spider 320 and the hydrokinetic device 330 is a gear train including a planet carrier 323 which is axially positionable along shaft 324 by the positioning fork 319 and clutchingly engages the spider 320 at 321. Planet carrier 323 carries the planetary pinions 325 which mesh with the gear 327 of the reversing cluster of planetary gears including the gears 327 and 329. Gears 329 mesh with the sun-gear 333 which drive the input torus 335 of hydrokinetic device 330. Disposed about this gear train is a ring-gear means 337 which is axially displacable by the positioning fork 339 and clutchingly engages a flange of input torus 335 at 341.

A brake band 343 is disposed about ring-gear 337 for selectively locking it against rotation relative to the transmission housing. The first torus 335 includes a flange 326 which forms a clutching surface 328 which is selectively engaged by a flange 350. Flange 350 is affixed to a sleeve 351 which carries a gear 346 at its opposite end and rotates co-axial with input shaft 324. Sleeve 351 is axially displacable along shaft 324 by a fourth positioning fork 342. The second torus 332 of the hydrokinetic device 330 is spline locked to sleeve 351 at 353.

A brake band 345 is disposed concentric with the first torus 335 for selectively looking it relative to housing 317 and a brake band 356 is provided concentric with flange 350 for selectively locking it relative to housing 317. An output shaft 364 is connected to the sun-gear 366 of an output planetary system 367. Sungear 366 is operatively coupled to input shaft 324 by three planetary pinions 368 that are carried by the spider 370 affixed to the end of input shaft 324. Planetary pinions 368 also engage a ring-gear 372 which drives a sleeve 378 that is journaled about input shaft 324 by a suitable bearing member. Affixed to sleeve 378 is an annular worm gear 388.

Rotatably mounted concentric with sleeve 378 and about worm gear 388 is a reactor frame 394 having an outer periphery forming a braking surface 396. A braking band 398 is provided for engaging surface 396 to selectively lock frame 394 relative to housing 317. Frame 394 carries a first set of control gears 400 and 402 similar to those illustrated in FIG. 10. As indicated, the gears 400 and 402 are mounted on a common shaft which is journaled to frame 394 by suitable bearing means. Gear 400 is a helical gear suited for operatively engaging worm gear 388.

Also carried by frame 394 is a second set of control gears 410 and 412 similar to the control gears shown in FIG. 11. These gears are mounted on a shaft 414 that is journaled to frame 394. Gear 410 meshes with gear 346 and gear 412 operatively engages the gear 402.

By actuating the various brake bands and clutching mechanisms in various combinations, this embodiment of the present invention can be made to assume at least nine operative drive configurations including a low gear configuration, two intermediate gear configurations, a direct drive configuration, and five overdrive configurations. In addition, neutral configurations are also obtainable as described in the previous embodiments. In the low gear configuration, lockout clutch 312 is disengaged so that flywheel 300 drives sun-gear 306, which in turn drives planetary pinions 308 around the inner surface of ring-gear 310 which is locked in place by the actuation of brake band 316. This of course causes spider 320 to rotate with the planetary pinions 308 driving planet carrier 323 via the clutching mechanism 321. In this configuration, clutching mechanisms 341 and 328 are also engaged so that all rotatable components to the left of planetary carrier 323 rotate as a unit thereby driving output shaft 364 at the low speed determined by the first planetary system 315.

The transmission is shifted into the intermediate gear configuration by disengaging clutch 321 and clutch 328, and engaging brake bands 356 and 398. In this configuration, the drive is transferred from flywheel 300 through the first ring-gear assembly 315 to shaft 324 which in turn drives the output planetary system 316. Since brake bands 356 and 398 are actuated, the interconnecting gear assemblies are locked thereby in turn locking ring-gear 372 against rotation so that input shaft 324 causes the planetary pinions 368 to be rotated about the inner periphery of ring-gear 372 thereby driving sun-gear 366 and output shaft 364.

When the speed of the vehicle has reached a suitable speed, conventional shifting mechanism causes the transmission to be shifted into the direct drive configuration by releasing brake bands 316, 356 and 398 engaging clutch 312, clutch 321 and clutch 328. In this configuration it will be appreciated that the first planetary system 315 rotates as a unit and in turn drives all rotatable components within the transmission as a unit so that there is, in fact, a true direct drive effected between flywheel 300 and output shaft 364.

When the speed of the vehicle is such that an overdrive configuration is appropriate, the first overdrive configuration may be effectuated by disengaging clutches 341 and 328 and actuating brake band 343. In this instance, the first planetary system 315 rotates as a unit causing spider 320 to be rotated directly with flywheel 300 to drive planet carrier 323. Since brake band 343 is actuated to prevent rotation of the ringgear 337, it will be noted that the planetary pinions 325 are caused to rotate around the inner periphery of ringgear 337 driving the planetary gear system including gears 327 and 329. These gears in turn are driven by the sun-gear 333 which is driven by (feeback) the first torus 335 of the hydrokinetic device 330. Since spider 320 is also affixed to input shaft 324, rotation of spider 370 will tend to cause sun-gear 366 and output shaft 364 to be rotated at a speed greater than that of shaft 324 which is directly related to the relationship between the drag imposed on ring-gear 372 and the resisting torque developed in output shaft 364. The drag applied to ring-gear 372 is applied by the second torus 332 of the hydrokinetic device 330 via the gear train comprised of worm wheel 388, gears 400, 402, 412, 410 and 346.

In the second overdrive configuration, brake band 398 is actuated to prevent frame 394 from rotating,

thereby increasing the speed of the second torus 332 of the fluid clutch and thereby increasing the resistance to rotation of the regulator members, the reduced speed of which results in increasing the speed of the output shaft.

In the third overdrive configuration, clutch 321 is disengaged and brake band 345 is actuated to prevent rotation of the first torus 335 of hydrokinetic device 330. In this configuration, the differential rotation between the first and second torus of hydrokinetic device 330 is increased thereby again increasing the drag applied to ring-gear 372. This results in further slowing down the rotation of the regulator members which are now rotating very slowly resulting in increased speed of the output shaft.

The maximum fixed ratio, overdrive condition is achieved by merely actuating brake band 356 so as to hold rotation of brake drum 350 and thus gear 346. With proportions of six for ring-gear 372 and 4.8 for sun-gear 366, a fixed overdrive ratio of l to 2 A is obtained, i.e., for each revolution of the crankshaft there will be 2 54 revolutions of the output shaft. If the reversing gearing and thefirst torus gearing are made separate assemblies with suitable controls, interacting brakes and clutches, the number of of configurations and possible design variations could be significantly increased.

As used in the claims, the term selectable ratio means is deemed to mean any of the various elements described above, or their equivalents, which permit the drag ratio of the reactor means to be selectively varied.

Whereas the present invention has been described above in terms of several specific embodiments, it will be appreciated that many alterations, modifications and/or combinations of the various components could be utilized for particular applications. Accordingly, it is intended that the several embodiments be recognized as exemplary with no aspect of limitation intended. Moreover, it is intended that the appended claims be interpreted as covering all alterations, modifications and/or combinations as fall within the true spirit and scope of the invention.

What is claimed is:

l. A power transmission apparatus, comprising:

an input shaft;

input drive means responsive to input energy and operative to impart rotation to said input shaft at selected drive ratios, said input drive means including an input planetary gear system having an input energy driven input sun gear, an input ring gear and a plurality of input planetary pinions disposed in engaging relationship with said ring gear and said input sun gear, said input pinions being driveably coupled to said input shaft, and braking means for selectively controlling rotation of said input ring gear;

an output shaft;

output drive means responsive to rotation of said input shaft and operative to impart rotation to said output shaft, said output drive means including an output planetary gear system having an output sun gear connected to said input shaft, an output ring gear and a plurality of output planetary pinions disposed in engaging relationship with said output ring gear and said output sun gear, said output planetary pinions being drivably coupled to said output shaft; and

reactor means including a gear train, clutching means and selectable ratio means driveably coupled in series between said output ring gear and said input shaft, whereby said reactor means enables a variable drag force to be transmitted to said output ring gear to cause the drive ratio between said input shaft and said output shaft to exceed 1:1.

2. A power transmission apparatus as recited in claim 1 wherein said clutching means includes a hydrokinetic device having an impeller element coupled to said input shaft by said selectable ratio means, and a runner 4. A power transmission apparatus as recited in claim 2 wherein said selectable ratio means includes a plurality of gears selectively interconnectable between said impellor element and said input shaft enabling the magnitude of the drag force applied to said output ring gear to be selected. 

1. A power transmission apparatus, comprising: an input shaft; input drive means responsive to input energy and operative to impart rotation to said input shaft at selected drive ratios, said input drive means including an input planetary gear system having an input energy driven input sun gear, an input ring gear and a plurality of input planeTary pinions disposed in engaging relationship with said ring gear and said input sun gear, said input pinions being driveably coupled to said input shaft, and braking means for selectively controlling rotation of said input ring gear; an output shaft; output drive means responsive to rotation of said input shaft and operative to impart rotation to said output shaft, said output drive means including an output planetary gear system having an output sun gear connected to said input shaft, an output ring gear and a plurality of output planetary pinions disposed in engaging relationship with said output ring gear and said output sun gear, said output planetary pinions being drivably coupled to said output shaft; and reactor means including a gear train, clutching means and selectable ratio means driveably coupled in series between said output ring gear and said input shaft, whereby said reactor means enables a variable drag force to be transmitted to said output ring gear to cause the drive ratio between said input shaft and said output shaft to exceed 1:1.
 2. A power transmission apparatus as recited in claim 1 wherein said clutching means includes a hydrokinetic device having an impeller element coupled to said input shaft by said selectable ratio means, and a runner element coupled to said output ring gear by said gear train.
 3. A power transmission apparatus as recited in claim 2 wherein said gear train includes a plurality of gears coupling said output ring gear to said runner element and is operative to transfer rotary energy from said output ring gear to said runner element in opposition to rotation of said impeller element.
 4. A power transmission apparatus as recited in claim 2 wherein said selectable ratio means includes a plurality of gears selectively interconnectable between said impellor element and said input shaft enabling the magnitude of the drag force applied to said output ring gear to be selected. 